1. Field of the Invention
This invention is directed to material and methods for reducing vibration in structures.
2. Description of the Related Art
Resonance occurs when a frequency is introduced at which a structure naturally vibrates. When a structure is excited at one or more of its resonant frequencies (for purposes of this invention, referred to as “natural frequencies”), an amplified motion (in terms of displacement, velocity, or acceleration) response occurs. When resonant excitation occurs within the audible frequency range (20 Hz to 20,000 Hz), the result is sound that, if of sufficient amplitude, is objectionable. Objectionable vibration may occur when a structure undergoes resonant response at any frequency.
Various active and passive noise reduction control techniques are known and used to control structural vibration and accompanying sound radiation. For purposes of this invention, vibration will be considered as including sound. Active vibration control systems use sensors to measure the amplitude and phase of vibration and/or noise from a vibrating structure. The sensed vibration or noise is inverted and fed to an actuator or loudspeaker to cancel the troublesome vibration or noise. In practice this technique reduces the vibration or noise significantly but does not eliminate it altogether. Active control systems are typically effective at lower frequencies such as below 1000 Hz. In many instances active noise reduction techniques adequately reduce vibrations and noise, but at the cost of expensive and complex sensing/actuation/feedback control/connectivity systems.
In contrast to active control systems, passive vibration and noise control systems made in sheet form usually are less complex and less costly. However, passive control systems can have significant mass and are typically practical only at frequencies above 500 Hz since it is at these relatively higher frequencies that the dimensions of the passive control systems are comparable with the relatively short wavelength of the vibration of the vibrating body. Application of a passive vibration or sound control system is generally ineffective as the physical thickness is generally too large and the mass is generally too high, therefore flexibility of such systems is limited to conform to non-flat systems, such as wall cavity, spatially challenging structures, pipes.
A third alternative is a passive control system known as a tuned vibration absorber (TVA). When using a tuned vibration absorber, a spring-mass system is tuned to vibrate at a frequency of interest, e.g. the same vibration frequency of the structure undergoing troublesome vibration onto which the vibration absorber is attached or mounted. In use, at the tuning frequency of interest, the tuned vibration absorber vibrates out-of-phase with the troublesome structural vibration and applies a force opposite the motion of the structure, thus reducing the original structure's motion response.
So-called point tuned vibration absorbers are an effective method of reducing the noise or vibration of a structure. However, a point absorber only controls the vibration or noise at one frequency at one point on the structure and is thus limited in its function to control vibrations over a large area of the vibrating body.
Most real structures vibrate at many frequencies simultaneously when excited by a broad band function (noise and/or vibration). For purposes of this invention, a structure is defined as “modally dense.” when the natural frequencies are closely spaced apart in the frequency domain, For example, major structural components used in buildings, such as single and double stud walls, floors, and ceilings, are modally dense vibrating structures as a result of the mass and geometry of the components, nonuniform properties of the building components, such as nonuniform density or thickness, the complicated joinery, attachment and mounting methods used in such building components and the resultant boundary conditions. Metal enclosures conventionally used for various kinds of industrial equipment, such as, for example, fans, chillers, motors, air handlers, pumps, generators, compressors, etc., are also modally dense vibrating structures.
There is a concern when applying a single degree of freedom tuned vibration absorber to modally dense structures. That is, when a single degree of freedom tuned vibration absorber tuned to a given natural frequency is applied to the structure that has excessive vibration or noise, the resonant response at the targeted natural frequency is reduced but the resonant response is increased at two new frequencies, one at a frequency lower than the targeted natural frequency (at which the absorber mass moves in phase with the structural mass) and another at a frequency higher than the targeted natural frequency (at which the absorber mass moves out-of-phase with the structural mass). That is, the targeted mode is split into two response frequencies, which in turn will be superimposed on the pre-existing structural modes, and will increase noise and/or vibration at one or both of the two response frequencies if the pre-existing response is in phase with the new modal response. In fact, the sound response (sound pressure level (SPL)) and/or the vibration response (root mean square (RMS) displacement, RMS velocity, or RMS acceleration) at the two new frequencies can be undesirably large, larger than the original frequency response at the two specific frequencies which was low before the vibration absorber was applied. Thus, in a modally dense vibrating structure having multiple closely spaced natural frequencies, the application of a single degree of freedom vibration absorber tuned to one natural frequency can result in an undesirable increase in vibration and/or acoustic radiation at adjacent natural frequencies.
In contrast to “modally dense” structures, “modally sparse” structures exhibit a frequency response function in which the majority of natural frequencies are not affected by the application of single degree of freedom vibration absorbers tuned to adjacent natural frequencies. Such structures are typically ideal or laboratory scale structures. FIGS. 1a and 1b illustrate the frequency distributions of modally sparse and modally dense structures, respectively. In the response function of FIG. 1a, the majority of resonant responses are not affected by neighboring modes. In contrast, in the response function of FIG. 1b, the majority of resonant responses overlap.
The spacing of the natural frequencies in a modally dense structure depends upon the frequency range of interest. Natural frequencies spaced about 40 Hz or less may be sufficient for a structure to be considered modally dense at lower frequencies, such as below about 500 Hz, whereas at higher frequency ranges, natural frequencies could be spaced less closely apart and still be considered modally dense.
None of the known passive vibration and acoustic absorbers adequately controls vibration and/or acoustic radiation from modally dense vibrating structures. It would be desirable to have a passive system which would reduce unwanted vibration and/or noise from a modally dense vibrating structure at multiple natural frequencies, particularly over a continuous frequency range, without introducing additional unwanted vibration and/or noise at other frequencies.